Compound power transmission



0. K, KELLEY 2A33 052 COMPOUND POWER TRANSMISSION Filed Sept. 6, 1943 5Shams-Sheet l 3mnemor Dec. 23, 1947. 0, KELLEY 2,433,052

COMPOUND POWER TRANSMISS ION Filed Sept. 6, 1943 5 Sheets-Sheet 2 Dec.23, 1 947. 0, KELLEY 2,433,052

COMPOUND POWER TRANSMISSICN Filed Sept. 6, 1943 5 Sheets-Sheet s SummerWyn/F161 pv i Dec. 23, 1947. 0. K. KELLEY COMPOUND P OWER TRANSMISSION 5Sheets-Sheet 4 Filed Sept. 6, 1943 Y Cali J79,

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I courounn [POWER TRARSIISSION Filed Sept. 6,1943 5, Sheets-Sheet 5Snnento:

Patented Dec. 23, 1947 2,433,052 COMPOUND POWER TRANSMISSIQN Oliver K.Kelley, Birmingham, Micln, assignor to General Motors Corporation,Detroit, Mich., a corporation of Delaware Application September 6, 1943,Serial No. 501,389

20 Claims.

The present invention relates to power transmissions for heavy dutypurposes, especially for those in which the engine drives throughtorqueconverting mechanisms and gearing always indirectly and withoutbeing positively coupled at unitary speed with the load shaft. Thismethod oi! drive is adapted to tractors, mining gear, oil-well drillingmachines, excavators, military vehicles and the like, and is-of especialreference to Power plant assemblies in which the engine torque can bestbe handled by coupling the engine through torque-converting means,stepping up the intermediate speed ratios into overdrive or overspeedranges, and stepping down to required final drive speeds throughtransfer gearing of differential or similar types.

In order to make full utility of the efficiency of a torque-converterhaving a definite, efllcient infinitely-variable speed ratio range, thepresent invention contemplates the use, in combination, of a powersource having relatively high. torque available at lower speed ranges,which may be a Diesel or steam engine; in operating that or other engineat or near its peak torque, while allowing the torque-converter drivenby the engine to deliver an output torque in accordance with theprevailing rotor load and speeds of the impeller; and changing the netspeed ratio between the torque converter and the load shaft with agearing device providing direct coupling, and two faster geared speeds.

With this regime it is possible to maintain the power plant at nearlyconstant speed, at approxlq mately peak torque, since in the drivebetween engine and load shaft, if there is no interruption in torque forratio shift purposes, the engine need never be throttled, after theloadshaft has been put in motion, except for stopping. The stall speed ofthe torque converter may be nearly equal to the maximum allowed speed ofthe engine, so that with proper design of the torque converter unit, ahighly efllcient use of engine and torque multiplier is made, the torqueconverter being always operated at speeds approaching and not exceedingits maximum torque capacity. In the example to be given in the presentspecification, the torque converter has a useful speed ratio range offrom one-to-six up to one-to-two, input to output. As shall bedemonstrated, this ratio range superimposedupon that of the gearbox inseries with the torque converter unit, provides a net ratio rangebetween engine and output of high flexibility and torque-handlingability.

The power-developing and controlling methods shown herein will appearunorthodox to'inventors 65 of conventional drives, but there aresubstantial reasons for the novel arrangement disclosed. The

- variable speed ratio over a range'of ratios at fairly high efficiency.but loses eiliciency as its speed ratio approaches 1 to l. Orthodoxreduction gear drives in series with torque converters of this type donot combine the eflicient ranges of such gear drives with the efficientrange of fluid torque converter drives; for lugging heavy loads atmaximum, peak engine torque, since all step-ratio gears are lessefllcient, proportionally to the de-' gree' of reduction, than when indirect-coupled drive. Furthermore, such reduction gear andtorque-converter combinations are not adaptable to dual power-plant,track laying drives, experience teaching that in reduction, epicyclicgears, having low differential or reaction brake factors,

the glaze and chatter effects are pronounced, and

excessive wear occurs, thus accelerating the service time and cost; and,in the case of the military vehicle, impairing the continuity ofutility.

The present invention therefore provides a high efilciency, low ratiofor drive with the fluid torque converter operating at best efliciencyfor pulling a heavy load, the engine being at maximum, or peak torque,and is especially adaptable because of its use of overspeed gearingbetween the engine-driven torque converter and the transfer,reductiongear driven by the gearing, since the reaction brake factor inan overspeed gear as described herein is high, therefore the risk ofchatter and torque shock is relatively nil. The invention also providessimple control means for actuating the ratio-establishing structures ofthe driving mechanism disclosed; and shows how these means are to beused for emergency braking ofthe vehicle. These and other advantageswill appear in the following specification describing the accompanyingdrawings, briefly characterized as follows:

Figure 1' is a; longitudinal elevation section through the transmissionassembly of the invention, the letters A, B and C denoting thecompartments in the casing for the torque converter, the gearing unitand the coupling clutches, respectively. Figure 1a is a section of aninput drive gear for the structure oi Figure 1.

Figure 2 is a part section at 2-4 of Figure 1, showing a typicalactuation control for one ofthe ratio-determining brakes of Figure l. Asimilar up of shell 4 to which view could be made for the second oiv thetwo brakes.

Figure 3 is a servo supply diagram 1 the actuation system for the brakesand clutches of Figure 1, with the valving and actuating cylinders shownschematically, and one pump unit with regulator valve for the pump linepressure.

Figures 4, 5 and 6 are provided to show the three ratio controlpositions of one of the valves of Figure 3, and its eflect upon theactuation unit controlled, the heavy arrows denoting active pump linepressure and the light dashed arrows, the relief pressure.

Figure 7 is a side-elevation view of a control valve box for the twovalves of Figure 3, with the mechanism for positioning the valves forestablishing the various of the assembly of Figure 1.

Figure 8 illustrates a typical vehicle control installation for one ofthe assemblies of Figure 1,

with levers, shafts and rods external to, and for shifting the mechanismof Figure 7 to operate the valves of Figure 3.

Figure 9 shows one of the assemblies of Figure 1 as it would be set upfor use in a vehicle, 'the valve box with its cover lifia housing theconstruction of Figure 7; and the final drive being through a reductiongearing driven from the transmission output shaft.

Figure 10 is a schematic showing of one of the ratio control valves ofFigure 3 equipped with a supplementary valve for compelling simultaneousbrake and clutch operation in the structure of Figure 1. The Figure 11view includes fragmentary parts showings taken from Figure 2.

Figure 11 is a schematic projection view of the control of Figure 8modified to operate the con- 1 struction of Figure 10, and includes afragmentary showing of some of the parts of Figure 7 in order to orientthe modification with the foregoing showings.

Figure 12 is a typical chart of the power results obtainable with thestructures herein described.

In Figure 1,'the input or power shaft I is connected directly orindirectly to an engine, by flange 2, supported in the casing I00 byappropriate bearings and is bolted to impeller unit 3 of thetorqueconverter. The impeller 3 is made a series of blades 5 is affixed. suported by ring 6.

The turbine output shaft l0, supported in the 50- casing' I90 by properbearings is splined to the turbine output member ll made up of shell 12with aifixed blades l3 and core ring it which in turn carries blades l5and 16. The output rotor therefore has three courses of blades [3, I 5and IS. A stator member of the non-rotating casing Hill has two sets ofblades l8 and I9 as reaction means for the turbine action oftransmitting multiplied torque.

In this device there are three transfers of torques. The spinningimpeller 3 imparts acomponent to blades I5. The movement of fluidthrough the reaction or guide blades l8 imparts a forward component toblades l6, and the reaction eifect of motion through blades l9 applies afurther component toblades l3.

By selected design of the parts, the available range oftorque-multiplying characteristics of the unit may be chosen to suit thticular vehicle to which the turbine device is fitted. In the presentinstallation. for reasons which will be later understood, a speed ratioreduction range through the torque converter unit providing a maximum of1 to 6 is desirable.

The turbine output shaft l0 extends to the) right in Figure 1. and isformed to support the spindles 2| of planet gears 22 of the overspeedand reverse gear unit, of double reduction type, the fianging andspindle support ring acting as a carrier 20. 1

The turbine unit is required to operate at a given filling of liquid,and seals 1 and 8 assist in maintaining the pressure level in theworking space. The unit may use the.oil from the transmission casing, oran independent supply, as desired; The unit shown herein is arranged tobe maintained under a static pressure head when it is delivering drivingtorque.

The fundamental principles of the fluid turbine torque converter areshown 'in Letters Patent to Fiittinger, U. S. 1,199,359, issuedSeptember 26, 1916, and are not deemed necessary of further descriptionin this specification. Other forms of continuously-variable torqueconverters may be used in my inventive combination.

The present invention combines the torque converter drive with gearing.in a way believed novel, and for definite purposes, as will be furtherdisclosed. This combination provides an unusual power plant output,especially for heavy vehicles in the drive of which the engine is neverconnected directly at unit speed with the output shaft.

The following description covers the overspeed and reverse drive gearingcombined with the fluid torque converter, by way of example.

The power input member of the double reduction gear unit is carrier 20of the primary group, which mesh internally with sun gear 23, andexternally with annulus gear 24. The transmission shaft 25 is fixed tosun gear 23. to sun gear 26 of the secondary group, and is attached tosplined clutch drum 36, upon which clutch plates 39 are slidinglymounted.

The carrier 27, for planets 28 of the secondary group is attached toannulus gear 24 of the primary group, and its drum 29 has an externalsurface upon which brake 80 may be clamped. Planets 28 mesh internallywith sun gear 26 and externally with annulus gear 3|, fixed or integralwith hollow shaft 32 of splined clutch drum 33,

, the external surface of the annulus gear 3| being brakable'by brake35. o

e needs of the par- The load, or output shaft 50 carries overhangingdrum 49, supporting the sets of clutch plates 81, and 38, mating withplates 3d and 39 respectively of drums 33 and 88..

There can be no drive between shafts l0 and Bil until an energisation ofa combination of at least one brake and one clutch or both clutches isaccomplished.

To initiate drive, both clutches tit-31 and 38.39 are energised, alocking couple being established across annulus gear 3| and sun gear 26,and across annulus gear 24 and sun gear 23. This provides 1-to-1 driveratio between shafts l0 and 50. To obtain the next highest speed ratio,clutch 243| is released and brake 35 applied to drum ll of stoppingrotation of annulus gear 3| oi the secondary group. The rotation ofshaft l0 and carrier 20 causes.

forwardly faster than the transmission shaft 25, cut forwardly onannulus of the secondary unit. The 21 reacts'from annulus gear 2 andshaft 25 forwardly, at

gear 24 and carrier 2'! component on carrier 9| to drive sun gear 2overspeed.

The peculiar ratio eflects obtainable by combinations of doublereduction gearing are not widefor the planet gears 22 planet gears 22 torotate 1y known, are somewhat difllcult to grasp, and in explainingclearly how the ratio effects herein are obtained, it is believed usefulto state something of the theory on which the unique drivemechafunctions of double reduction planetary gear units. To estimatethese factors, it is required first, to figure the groups separately.and to superimpose from one to the other, the added characteristicderived from rotation of the secondary upon the primary.

For example, in the Figure 1 presentation, it being desired to know thespeed ratio of shaft 25 to shaft l0, arbitrary values are first assignedthe different gear elements, and the net ratios of the unit thencalculated.

The pitch diameters or numbers of teeth of the elements may be used toobtain the ratios, in a commonly known manner.

Assuming the sun gears 23 and 26 to have the same diameters, the annulusgears 24 and 3| to have the same diameters, by assigning arbitrary toothor diameter values of 2.0 to the sun gears and 6.0 to the annulus gears,the ratio between shafts 25 and I0 may-be simply calculated by turningshaft 25 one turn, and deriving the fundamental component applied toshaft l0 through the primary gear group, with sun gear 23 as the driver,and annulus gear 24 held. This component will be a value represented bythe sun gear diameter divided by the sum of the annulus and ring geardiameters, or 2+8, which equals 0.25.

The added component derived from rotation of annulus gear 24 because ofit being. attached to carrier 21 of the secondary group is found byfiguring the rotation effect on carrier with one turn given to annulusgear 24 with sun gear 23 held, and multiplying that by the actualcomponent applied to carrier 21 through rotation of sun gear 26 of thesecondary group, reacting from the annulus gear 3! held againstrotation.

One turn of annulus 24 with sun gear 23 temporarily held, in the primarygroup, would give a value equal to the annulusdiameter divided by thesum of the annulus and sun gear diameters, or would be equal to 6+8 orthree quarters. 0.75. But since the annulus does not make one turn, butmakes a fraction thereof, conditioned by the interaction within thesecondary group, identical with that first given which yielded 2+8 or0.25,

the incremental component applied to the annulus gear 24 of the primarygroup, derived from the secondary group is one-fourth of threefourths orthree-sixteenths, which is to be added to the primary component ofone-fourth ,to achieve the net rotation of carrier 20 and shaft l0 for lrotation of shaft 25 with annulus gear 3| of the secondary group held bybrake 35.

The net rotation of shaft I0 is therefore sevensixteenths for one turnof shaft 25, or 0.4375. Conversely for one turn of shaft l0, shaft 25 isturned at overspeed 2.286 times. As explained further, this ratio isOverspeed-l.

Now with a simple planetary gear interaction in the primary group, withbrake held only, and brake not held, the secondary group takes nopart.in the ratio determining action at all.

From the above, we found that with a simple planetary relationship, oneturn of shaft 25 gave one-quarter turn to shaft I0, so conversely, forone turn of shaft In the shaft-25 would rotate at overspeed 4.0 times.

Thusby merely shifting actuation from brake 30 to brake 35. whileretaining clutch 38-38 encaged, the shaft 25 can be driven at 2.286overspeed ratio. and be speeded up to an overspeed ratio of 4.0. Whenthis occurs. since annulus 24 and carrier 21 a e held from rotation, theidling annulus gear 3| and attached shaft 22, clutch drum 33 and plates34 spin backward freely. This ratio'is overspeed-2.

Downshift" to first overspeed from second overspeed is as easilyaccomplished by releasing brake 30 and actuating brake 35, and return todirect coupled drive may be then obtained by releasing brake 35 andengaging clutch 34-31.

In setting up the controls for this power transmission, it is notnecessary to declutch one or both of the clutches 34-31 or 38--38. Itshould be chserved that with either one of the clutches disengaged, notorque can flow while both brakes are released. even if the engine isidling fast enough to apply a torque to shaft 10. There can be no torquecomponent applied to shaft unless two of the brake or clutch members areengaged or locked.

This unusual arrangement is of utility, in that the controls, forexample, may be set up so that during direct, reverse or neutral, clutch34-31 need not be declutched at all, therefore simplifying the ratioselection pattern.

The following ratio pattern, as an example, will make the abovestatements clearly understandable, the symbol indicating units actuated:

Clutch Brake Brake Oversee d (1) overspeed (2) Not necessary to releasefor neutral.

taimng drive by merely increasing engine speed with the acceleratorpedal to a torque value capable of overcoming the load. Under extremeoperating conditions an immediate shift from Overspeed-2 to Reverse maybe accomplished by release of clutch 38-39 and actuation of clutch 3431.

For the transition from Direct to Overspeed--l; in the prior condition,all transmission elements are rotating at unit speed, brakes 30 and 35being released. The clutch 34-41 is then released and brake 35 applied.The braked parts 3l-3233 come to rest, and the shaft 25 and sun gears23, 26 increase speed ratio to 2.286 faster than shaft Hi, theintermediate gear elements also speeding up.

Drive in Overspeed-l may continue indefinitely. It must be rememberedthat the fluid turbine unit may be operating over a ratio range of 0.167to 0.5, so that the netoverallratio rangebetween shafts I and 25 mayvary in Overspeed-e-l from 0.38 to 1.143.

Shift to overspeed-2 requires release of brake 35 and actuation of brake30. The annulus gear 24 and carrier 21 must be brought to rest whileannulus gear 3|, shaft 32 and drum 33 must be started into motion andaccelerated reversely at some ratio to that imparted to sun gears 23 and26 and to shaft 25. Conversely,

7 downshift to verspeed1 requires resetting of brake 35, with release ofbrake l0, speeding up elements 24, 2! while bringing BI, 32, 33 to rest.

It will be noted that considerable inertia values exist on the partswhich idle during and after these transitions, so that the inertiaabsorption needed to complete a ratio transition may require a timelagto avoid building up high shock stresses on the shaftlng and criticalload points. It is therefore desirable to provide inertia absorptionmeans in the brakes 30 and 35 operative over suflicient time to avoidsuch stressing, and further, to so phase the actuation interval for onewith respect to the release of another, that a constant minimum torquevalue is maintained during the ratio transitions, for the purpose ofpreventing an idle rotating element from running free, to reduce thepossibility ing, and maintain efilcient drive torque converter.-

The prior art discloses control systems adapted for this purpose inwhich fluid servo actuation pressure means are arranged to .yield torqueoverlap, during the ratio shift interval, by beginning the applicationof pressure to one clutch or brake selected to assume the drive orreaction, before the pressure is relieved from that one selected torelease the drive or reaction.

The Letters Patent U. S. 2,193,304 to E. A. Thompson, issued March 12,1940, describes this type of control, and also'Letters Patent U. S.

through the of shock load-V sages I04, I05

2,204,872 to E. A. Thompson, issued June 18, 1940,-

it being obvious that the systems of control shown in these patents maybe adapted in part to operate the structure of the present invention.

To obtain reverse speed ratio, as noted above, it is necessary to lockbrake 30, free clutch 38-39 from drive, and transmit the drive to shaftfrom annulus gear 3| through clutch 343'I. Locking of brake 30 stopsrotation of annulus gear 24 and carrier 21. Rotation of carrier 20, withthe arbitrary dimensions used herein by way of example, causes planetgears to roll around the inside 'of annulus 24, imparting an overspeedto sun gear 23, the value of which, for one turn of carrier 20, is equalto 6.0 plus 2.0 divided by 2.0, or 4.0, as noted the explanation ofOverspeed-2. .This rotation being imparted to sun gear 26, withthe-carrier 21 held, rotates annulus gear 3| backward at a value equalto 2+6 for one turn of 26, or to one-third times 4, or 1 and-one-thirdturns of shaft 50 in reverse for one turn of shaft Ill.

The utilization of the disc clutches located at the delivery, end of thetransmission assembly remote from the fluid turbine torque converterprovides a. neat, close-coupled power transmission unit, capable ofready adaptation to special purpose vehicles of all types. military orotherwise.

The output shaft 50 is supported in casing I00 by a large double-rowbearing, and carries splined pinion which will gear with larger diameterdifferential or other-transfer gearing for applying the power planttorque to the final drive, wheels, tractor bull rings and the like.

As will be understood from the ratio shift pattern given preceding, itis requiredthat the actuation controls be able to load both clutches34-31 and 38 -39, and actuate them individually with the other released.

preceding in connection with,

. gear 63 is attached at 64 to Figure 2 is a part section taken at 22 oiFigure 1, for showing the method of actuation of both of the brakes 30or 35, the Figure 2 section referring to the actuation for brake 30. Thecas- 8 I00 is tapped and threaded to accommodate adjustable brake anchorbolt 52. The movable end of brake 30 is pivoted to thrust rod 53 seatingin notch 54a of bellcrank lever 55 pivoted on the casing, the long endof piston rod I55, because of the tension of brake release spring 53a.

Brake cylinder is fitted with piston 58 and rod 05, the pipes 91 and 95sure to either end of the cylinder 60 for applying or releasing thebrake 30. Biasing spring BI always tends to release the band 30,similarly to spring 53a, and the two springs must be overcome by thefluid pressure entering pipe 95, to shift piston 56- to the left.

This arrangement is duplicated for brake 30, the pressure piping beingshown further in Figure 3 in conjunction with that for the clutches.

Clutch 38-49 of Figure 1 isloaded by plate 42 guided on bolts 4|similarly to plates 38, and is shifted axially to cram the plates 38, 39against backing plate 43. Plate 42 is shifted to the left by pressurefed through transverse tubes IOI connecting space I02 with radialpassage I 03 and pasfed from gland space I08 and leading to externalpassage H0 in web I00c oi casing I00, as shown diagrammatically inFigure 3.

Clutch 3431 of Figure 1 is loaded by plate H2 uided on -bolts 4! of drum49, stud H3 guided in an aperture of web 49 of the drum, and flangedmember I I4 shiftable by fluid pressure fed into the spacer I I5 betweenweb 49b and member II l from passage I I6, gland passages I I7 and I I8,and external passage I20 in web I00c of casing I00.

Member H4 and plate 42 act as annular pistons, in loading the clutches34-31 and 38-39.

The location of web IOIlc between gear compartment B and clutchcompartment 0 provides a long bearing support against shaft deflectionand misalignment, as well as a convenient means for introducing theclutch actuating pressure.

The clutches are of to those described in Letters Patent to O. K.Kelley, U. 8. 2,176,138, issued October 17, 1939, and to E. A. Thompson,U. S. 2,193,524, issued March 12, 1940. These clutchesare self-released,and assisted in releasing by conventional coil springs which are seatedin the pressure plates, as is common in this art.

In Figure 1a is shown a section of the input gear which may be used forproviding a permanent multiplication of torque between the engine andthe torque converter input shaft.

An extension of the casing I00 is attached to the reaction sun gear 63of a flxed planetary unit.

The torque converter input shaft I has its flange 2 bolted to flange 2aof a jackshaft flanged at 62 to support the spindles of planet gears 59meshing with sun gear 63 and annulus gear 58, the latter being integralwith drum 5! bolted to the flywheel 9b and crankshaft end So. Theassembly is piloted at bearing 61, and the sleeve of sun the web I009.Lubricant is supplied through pipe 66 to nozzle as For this purpose,fluid pressure is admitted to shown, leakage being stopped by seal 68surrounding the shaft of carrier 62.

The external control passages H0 and I20 are fed by the servo pumpsthrough valving operated in accordance with the ratio shift patterngiven preceding.

which bears against feeding actuation presmultiple disc form, similar 9The following actuating pressure requirement for these passages willmake clear the method of The symbol P above indicates actuation pressureexisting. It may or may notbe desirable to maintain pressure onclutch-34-31 when the control is set for neutral." p

The assembly of driveelements of Figure 1 is accomplished bycompartmenting. 1 The torque converter unit occupies compartment A,having web portions of casing 193, as lllila. and I091) at the front andrear. Web N with web I00?) sets off the double-overspeed gearcompartment B, and the clutch compartment C is defined between webslllllc and 190d. There is no sump in the compartment A, nor incompartment 0. Excess oilin the latter, finds it way back to'the sumpwile through drain H30). Seal 9 at the rear of the assembly preventsescape of liquid along the shaft 53 from compartment C.

\ aasacsa right end of brake cylinder 89, of Figure 2, where the piston58 is actuated to oppose. spring it, which normally holds the brakeoperating, mechanism inactive. Therefore valve 10 which controls clutch34-31 also controls brake 35. For "Overspeed-2 however, it is requiredto hold both brake 35 and clutch 34-31 inactive, therefore the valve 10has this position in which it delivers pressure to brake cylinder toonly.

The gear pump indicated at 93 in Figure l maintains a head in the torqueconverter, over and above filling pressure, and for lubricating thedrive assembly bearings. The pump 98- may be used, likewise, to furnishfluid pressure for actuating the servo system' operating the describedbrakes and clutches.

Further, gear 5i cut on the periphery of the flange joining shaft W withdrum 49 may be used to drive a second'pump, not shown,'for taking overthe work of the pump 90, in whole or in part, when the vehicle is inmotion. The bottom space Hills of the casing we of Figure 1 is used asasump, and may be connected to the input of pump 36 in any convenientmanner, by piping or passages in the casing. The gear 5! is shownmeshing with gear tla spindled on shaft bib supported in web WM or thecasing, the latter gear being available to drive a second pump and agovernor shaft, as desired.

The open piping at the lower left'portion of Figure'B, as an extensionof the main pump line or of pump 9t may be fed from the second pumpdriven by shaft 50.

The ratio requirements call for the arrangement of Figure 2, utilizedfor controlling clutch 34-31 and brake 35 from a single valve 10, whichhas three positions; one, for actuating the clutch while holding off thebrake, another for releasing the clutch while permitting the brake toact, and another in which it holds off the brake.

The flow of pressure from pump line 92 in Figure 3 is blocked by thecentral boss 01' valve 19, when the valve is in the mid-position. Withno pressure availableto actuate either the brake 35 or the clutch 34-31.spring 6| holds off the brake 35 positively, and the Overspeed-2condition is fulfilled. With the valve 19 in the "up position, as shownin Figure 5. pump pressure may flow through port 1| and to lines 91 andHi] to energise the clutch 34-31 and to hold off the brake 35positively, the liquid trapped in the rightward portion of cylinder 6!!being relieved through line 95 and exhaust port 13. This is theoperating condition corresponding to Reverse, Neutral or Direct, in thepreceding ratio selection pattern.

Now with valve lfl in the down position the.

pump servo port H is open to port 14 and line 95, to provide pressure onthe rear of piston 56', acting agalnstspring 3|, to energise brake 35,corresponding to the required Overspeed-l condition.

The valve 10 is therefore held by the external controls in the upposition for establishing reverse, neutral and direct; is shifted to thedown position for Overspeed-l and to the mid" position for Overspeed-Z.

' Referring back to the above ratio selection chart, it will be notedthat a second valve'to control the brake 39 and the clutch 33-39 willneed to be able to establish actuation of brake 30 by itself, forReverse, release of both the brake and The arrangement of servo controland supply of Figure 3 consists of a pump 90, and regulator valve 3iported to supply a main servo line 92, having feeder leads 93 and 94 toinput ports 1i and M of valves ill and 80 respectively. The valves arebalanced valves having three positions, "up for delivery of the pressureto lines 91 or 98 for causing release of the. brakes 3t and '35 andclutch for"Neutral, and thereafter actuation of the clutch for the otherthree forward drive conditions.

The valve 80, identical with valve Ill, is used for controlling brake 30and clutch 33-39; in the down position for Reverse, in the 'mid-positionfor NeutraP and in the "up position for all forward driving.

actuation of the clutches 34-31 or 38-39;

down for relief of pressure via ports 12 and t3 while energising thebrakes 30 and 35, and a mid-position in which pressure drains from lines9], lit and 93, I23; and 95, 95. The upward protruding stems of thevalves 19 and 86 may be operated singly or together in a ratio selectionpattern as provided above.

The prescribed movements have to do with the actuation control ofclutches 34-31 and 38-39, and actuation or release of the brakes 39 or35. Brake 35 is never actuated when clutch 34-31 is v actuated,therefore the fluid pressure which loads this-clutch may be used to holdoff brake 35, the pressure feed in line 93 going to line 95 and to theThe ratio shift pattern given below will provide a clear summary of therequired valve positionings:

Valve ValveSG Reverse up down Neutr up l. m D1rect up"... up.Overspeed-l down up Overspccd-2 m up.

The three companion Figures 4, 5 and 6 illustrate the flow controlpattern of valve 10 of Figure 3, the Figure 4 showing duplicating theposition of that valve in the assembly diagram of Figure 3. In Figuredthe valve is blocking the pump line pressure, when neither the clutch orbrake are actuated. In Figure 5 the pump pressure holds the brake ofiwhile loading the clutch, and

. I24 and in Figure 6 the brake is applied by the pressure V while theclutch pressure is relieved.

The exhaust ports and lines relieve the spent pressure back to the sumpof compartment 3 whence the fluid is recirculated by pump 80 and anotherpump, not shown.

. Since the actuation mechanism for brake 30 is Identical with that forbrake 08. the diagram of Figure 3 for the brake cylinder, piston andother parts has them marked with prime members, 80' being the cylinder,58' the piston, 88' the piston rod, and 8| the release spring. Theoperation is the same for both. A duplicate showing for brake 35 likeFigure 2 is not believed necessar External controls "A simplevalve-setting means'for obtaining the above shift pattern is shown inFigure 7, where a cam plate I2I rotating on shaft I22 is slotted toaccommodate the pins I23 and I25 of levers I24 and I28 mounted on shaftI2? supported on valve casing I50. The slots I30 and I 3i are shaped toprovide the S-positional action required, and plate I2I is rotated byarm rod I33 operated by the structure of Figure 8.

The valves ,10 and 80 are shown protruding from the upper portion of thevalve casing G50, and clevises I24 and I28 of the arms of the levers I28provide positive positioning means for the three required positions ofeach valve.

The slot I30 of plate I 2I is cut so that the intersecting pin I23 oflever I24, controlling valve 10 for brake 35 and clutch 34-31, is heldclose to the centerline of shaft I22 for the "up" position of the valve,and farthest from the centerline for the down position.

The radial lines marked R," N," D," -1" and "0-2 Fig. 8 correspond tothe Reverse, Neutral, Direct, Overdrive-1" drivepositions of the valves.

The steering column I40 of the vehicle shown in Figure 8 supports shaftI II on which arm I42 pivoted to rod I83 is fixed, the manual handlevI48 attached to shaft I4I moving over sector plate I48 in accordancewith the selected ratio or drive condition. The handle and sectormechanism may be made in accordance with Figures 11, 12 and 13 ofLetters Patent U. S. 2,195,605 to E. A. Thompson, issued April'2, 1940.The leader action of the handle I45 of Figure 8 determines the followeraction of plate I2I which positions the valves 10 and 80 for therequired shift pattern.

The valve casing I50 in Figure 7 has a raised boss to act as a bearingfor the valve rock shaft.

I21, on which levers I24 and I28 pivot. These levers may be made ofinexpensive stamped parts,

and also plate I2I.

Various automatic methods of may be utilized, in which speed governorsconditioned by operator throttle setting or by torque, on a dynamometricbasis, may be used to shift the valving for establishing speed ratio inthe arrangement of the invention.

Figure 9 is an outline perspective view of the transmission assembly forthe purpose of showing the general relationships between the externalportions of the mechanism. The valve casing I50a is bolted to the maincasing I 00 as shown, and supports control shaft I22 and lever I32,shifted by rod I33, and the mechanism of Figure 7. Snap plates I and I52seal the casing space adjacent the brake operating mechanisms like thatof Figure 2, and are readily removed for inspection and adjustment-work,

The arrangement of Figure 9 may be placed I32 and pivoted a providingthree ranges of and "Over- 0f 3 to 1, or 3.3 to

selection of ratio as required in a standard or special Purpose vehicle.with the centerline of the drive assembly coincident with that of thevehicle, parallel to it at one side, transversely placed, or inwhatever-position needed for the particular drive, the linkage of Figure8 being arranged to coincide for the desired shift control throughproper bellcranks-and levers as called for by the adopted arrangement.It is not deemed necessary to show the final output drive other thanindicated schematically in Figure 9, where the part perspecthe viewshows the bevel gear 41 of shaft 50 meshed with larger gear 48 drivingcross shaft 48 at a reduction speed to that of shaft 50, for deliveringa powerful torque to shaft 48. It should be noted that the overallcombustion of an engine, a reduction speed ratio fluid torque converter,a multi-speed forward increasing and reverse gear and a step-down gearis believed novel for the purpose of driving a large and heavy vehicle.This sequence of units provides a constantly coupled torque multiplieryielding infinitely variable torque over a useful speed range in whichthe engine may be constantly operated at its torque peak, with aselective torque multiplication gear unit for directly delivering themultiplied torque of the converter, or delivering a 'reduced torque atincreased speed to an output shaft, finally reduced to a re-multipliedtorque to the final driving gear. The sequence is:

1. Engine at maximum torque,

2. Torque converter operating from about 6 to 1 torque multiplication toabout 2 to l, continuously variable;

3. Direct and two overspeeds forward, selective:

torque to output, infinitely variable because of the nature of thetorque converter,

4. Step down gear multiplying the de-multiplied torque of the gear unit,so that overall drive reduction suitable for heavy vehicles is obtained.

The final step-down gear of Figure 9 is not for the customary slow speedgear function, but is for the purpose of compensating for the overspeedgearing effect, since it is desirable with present-day engines to obtainan overall reduction drive range inheavy vehicles of 20 to 1, orthereabouts.

The overall reduction of 20 to 1 in the present invention isobtainedwith the gear unit in Direct, the torque converter operating atthe lower point of its curve, and with a final gear reduction 1, easilyobtainable with bevel gearing, as shown in Figures 1 and 9.

To endeavor to utilize a fluid torque converter and a common reductiongear drive unit in series with it would place a severe load factor onconventional gears, would not provide direct drive with no gear losseswhen the vehicle power would be called upon for sustaining maximumloads; and would drive or track-laying vehicles, due to the fact thatplanetary or epicyclie brakes so used, glaze and chatter at lowdifferential reaction speeds, while these phenomena do not occur athigher differentialreaction speeds. In the present invention, theunorthodox arrangement of step- ,down, step-up and step-down transfermeans allows the torque converter unit to, operate at its bestefiiciency with the gearbox at its best efilciency; i. e., when there isno relative gear rotation action;

tial reaction speed factor on the overdrive bands 30 and 35, whichoperate to set up all of the forward drive ratio shifts beyond initialdrive be diflicultly adaptable to dualand it provides a high difleren-13' i when clutches 34-31 and 38-39 are first actuated.

In the foregoing description, the novel features shown include a form ofdrive through-a fluid turbine torque converter, in which intermediatespeed ratios are stepped up into overspeed ranges and stepped down tofinal drive ratios; 9. combination inwhich a high-torque power sourcemay be operated advantageously at its power peak, while allowing theconverter to deliver its output torque in accordance with prevailingrotor load and impeller speeds; and a combination in which the net speedratio between the converter and the output shaft is changed with agearing device which provides direct coupling and two faster gearedspeed ratios.

An advantage lies in the ability of this mechanism to maintain the powerplant at its peak power, and at constant speed, in that there is nointerruption of torque for speed ratio changing and consequently node-throttling, except for stopping and starting. A further advantagelies in the fact that the design stall speed of the converter is madenearly equal to the maximum allowed engine speed, providing a match ofpeak emciencies within a small speed range for both engine andconverter, with the safety feature of having the combination approach,but not exceed the maximum torque capacity of the converter, whichavoids the excessive churning losses of heat'in the converter fluid.

As will be readily understood by one skilled in the art, theseadvantages are attained on a flexible operation basis, that is, theco-action efiect between power plant and converter is not made rigid forjust a small range of speed, but is likewise available for throttlesettings and speeds under those for which the power peak and stall speedrequirement are established. The period of time during which, forvarying drive conditions, the effective and eflicient use of thesequalities is attained, is thereby lengthened, and consequently theoverall utility of the torque converter, without undue heat-loss is muchgreater than that of any known drive system for heavy, constant-pullpurpose vehicle drives.

fore described. The movement of the handle I"- undesirable passingthrough of a position such as valve 13 which upon shift between "Directand 'Overspeed-l must pass through its mid-position in which it blocksoff the pump line pressure and would drain brake cylinder 60 and thepressure compartment of clutch 34-31; Too long a dwell of handle 5-betwee D and "0-1" would therefore, on the upshift, release clutch 34-31without actuating brake 35 so that a nodrive interval would occur. Toolong a dwell on the downshift would conversely release brake 35 withoutactuating clutch 34-31; giving a similar interval, so that undesirableracing of the engine connected parts could occur in either case,followed by a surge of torque as the brake or clutch to be next actuatedon the transition, would be loaded. The inclusion of poppeting of thehandle M5 for snapping over between D and "0-1." positions, obviatesthis effect, only present for the noted shift, when valve Hi makes afull stroke between its extreme positions.

The controls as described in Figure '7 do not provide for brakes an andas to be simultaneously In view of the rather full description of the"operation of each of the portions and units of the invention givenabove, it is not believed necessary to provide a detailed description ofthe full operation sequence, ratio to ratio. Assuming the engine drivingthe shaft l of Figure 1 idling with the vehicle stopped, the impeller ofthe torque converter will not provide a driving torque in that unituntil the engine speed is advanced sufficiently. With the operatorshandle I of Fi ure 8 in N position on the sector I45, the valves ill andBil are positioned to hold both brakes 30 and 35 from actuation, andneither clutch 36-31 or 38-39 is engaged. Movement of handle 5 to Dpositions the val'ving to cause actuation of both clutches 34-31 and38-39, the brakes 30 and 35 being inactive, and shaft 25 is driven atthe same speed as shaft ill. Shifting of the hanclle to O-1"-re1easesclutch 34-31 and causes actuation of brake 35, the shaft 25 being-drivenat an overspeed ratio with respect to shaft l0.

Further shifting ,of the operators handle to "0-42" on the sector I45,positions the valvin'g to release brake 35 and to actuate brake 33, theeffect of which causes shaft 25 to be driven at a higher overspeed tothat of shaft I0. For reverse, the handle 5 in R position establishesactuation of brake 30 and clutch 34-31 with brake 35 and clutch 38-39released, as hereinbeloaded at any ratio point, since overload of thegearing and shafting would be experienced if clutches 3l-3'I or 38-39were engaged when this occurred. This point will be clear uponvinspection of. Figure 1. With clutch 34-31 only engaged, the braking ofannulus 3| by brake 35 would apply a, braking force to shaft '32 andthrough the clutch 34-31 to drum 49 and shaft 50, to stop the motion ofthe vehicle: the braking of brake 30 on drum 2! would establish a lockng couple across sun gears 23, 26 to carrier 20, thus slowing down rotorll of' the torque converter. With clutch 38-39 only engaged, the lockingof brakes 30 and 35 would similarly set up a locking 39 through theclutch. Assuming that in an emergency, it would be desirable to usebrakes 30 and 35 for slowing the shaft 50, and that the drive parts hadbeen designed to handle the torque loads thus created, it is clear thatthe use of both clutches 34-31 and 38-39 actuated together would avoidoverloading of one, by distributing the braking torque on the two,therefore a control such as diagrammed in Figure 10 provides a positiono the sector for handle M5 in which the brakes may be simultaneouslyoperated.

The Figures 3 to 6 demonstration provides that each valve 10 or B0divide the clutch actuation interval from the brake actuation interval.The

Figure 11 arrangement with the Figure 10 control clses all of thecontrol functions stated above for those valves. The supplementaryvehicle braking control valve 18 is utilized only when avehicle brakeeffect is desired, and is inactive and non-interfering at all othertimes. In an installation of this control the-construction of Figure 3would be replaced by two arrangements like Figure 10, one for each ofvalves Ill and 80, making two ratio control valves, and twosupplementary brake valves auxiliary thereto, so that all four elements.brakes II and 35 and clutches 34-31 and "-48 may be simultaneouslyactuated.

The supplementary braking valve 18 shown in Figure 10 compels actuationof the brake while the corresponding clutch is loaded for driving. Thepassage 81 does not connect directly with the left end of the brakecylinder 50a as in Figure 3, but is fed by passage 81a to port 15 ofbalanced valve 18. and from port 11 to the left end of the cylinder,through passage 55 instead of directly connecting the right end of thecylinder with the feed port 14. is connected through ports I8 and 19a tothe right end of the cylinder.

As shown in Figure 10, valve 18 causes no interruption in the normaloperation of valve 81 such as described for valve 15 of Figure 3, all ofthe control and actuation shifts for ratio change being the same.However. when valve 18 is raised, the feed from 91a to line 81b is shutoff, and its pressure is delivered to port 19a and line 85a to the rightend of cylinder 50a, to force piston 55a to the left and actuate thebrake. The servo pressure in line 91 therefore actuates the clutchthrough line I I and also the brake through passage 91a, ports 15 and19a and line 55a. As valve 13 raises, its second boss from the top shutsof! port 19a from line 55 so that the exhaust relief ordinarily presentwhen the clutch is actuated is now shut oil. Valve 15 has therefore twooperative positions, normally ,down in non-interfering position, and up"for simultaneous brake and clutch operation; when valve 81 is put inclutch actuation position.

The stem18a of valve 18 protrudes through the top of the valve casingI50,.threaded guide 181; serving to retain spring 180 and normallyholding the valve 18 in inactive position. The valve must be pulled "upagainst the action of spring 180, to compel simultaneous brake andclutch operation. I I

Since the clutch feed line IIII does not have pump pressure until thevalve 81 is in the "up position, and'since it is desirable to have bothclutches 34-31 and 3839 engaged when brakes 30 and 35 are energised, thevehicle braking control for this operation is best operated with theFigure 8 arrangement only when the drivers handie I45 is in the Dposition; that is, with both I clutches already actuated.

To avoid trapping fluid in the left end of cylinder Gilaand in passage911), which would oppose the application of the brake under pressurefrom passage 95a to the right end of cylinder 60a. the exhaust port I12ais provided to vent when the valve 18 is moved up.

To restrict operator abuse of thecontrol, the handle I45 may be adaptedto operate the stem of brake valve 18, the button I46 of Figure 11 beingpermitted to move inward flush with the head of the handle, at the D"position, in which it causes the valve 18 of Figure 11 to rise againstthe action of spring 18c. I

In Figure 11 a modified form of handle I45 is shown, for rocking shaftI4I, different from the Figure 8 showing. The handle I45.is hollow toaccommodate rod I41. and attached button I44, and is apertured at theknob for the inner portion of the button and to seat return spring I 48The crown head of the handle I45 adjacent the shaft houses the clevisend of rod I41 which is guided in the handle, and the short lever I49ais pivoted to the clevis, and has a swivel seat for the upper end of rodI49 guided in a hole drilled through the shaft I. The lower end of therod I49 emerges from the end of the shaft I and is bent to form a neckfor ball head I54, fitting cup I55 of lever I51 pivoted on the framing.The

other leg of lever I51 is pivoted to rod'I5B, which is in turn pivotedto lever I58 of shaft I21, the inner part of shaft I21 being fixed tolever I6! shown as clevised to the upper end of valve 18 of Figure 11.whose spring 130 normally holds the valve 18 in the down position. Theforce of spring 130 therefore tends to hold rod I58 to the right; rodI49 at its lowermost position, and therefore to augment the returnaction of spring I48 to keep button I44 in its outermost position withrespect to the knob of the handle. A second connection from lever I51 toa rod and lever similar to rod I58 and lever I59 may be used to operatea companion valve to the ratio valve 80, the schematic view of Figure 11only showing the linkage for the companion valve to valve 10.

When button I 44 is thrust inward, rod I49 is ,raised by the clevisaction, lever I51 rocks clockwise, rod I58 shifts to the left, leversI58 and I8I also rock clockwise, and valve 18 is lifted, which, fromreference to Figure 11, causes the brake 35 and clutch 34-31 to besimultaneously actuated.

In order to assure that this action occurs only when the handle I45 isin the D" position, a tapered lug I43 fastened to rod I'4'I protrudesthrough a window cut in the upper portion of the handle, and intersectsa tapered slot I 46a cut in the underside of the sector plate I46 at thecenter of the D" position. Slight oil-register of the handle iscorrected by the taper of lug I43 and slot I46a.

This arrangement is a convenience for the operator, since while itprovides braking action locked out against wrong action being required.

The chart of Figure 12 is to show the relative output torque values fordifferent operating conditions'and road speeds for a typicalinstallatlOll. 1

The scale at the left margin indicates the relative efllciency in thethree forward driving ratios, and the T T and I curves, thecorresponding output torques, while the upper black line curv representsthe relative variations in engine torque. It will be observed that therecommended point of shift from low forward or direct, to second forwardspeed ratio comes at a. point on the E curve at which the efficiencyfalls below percent. Therecommen'ded point for upshift to 3rd, orhighest ratio 0-2, is at about 24 M. P. H., where the E curve has begunto fall oil to about '72 percent. The drop in engine speed and torque atthese shift intervals is quite small, since the engine inertiadifferential is absorbed in the torque converter, being less at theshift point for upshift to 3rd, than from direct to 2nd. The scale atthe left may also be taken as proportional to tractive effort, as isknown in this art. The values given in this chart are approximatelycorrect for one specific installation, but

of course will vary with different vehicles, engines and required speedfor a given load.

l7 the torque converter characteristics in any we since the automaticreadjustment of speed ratio within the torque converter occurs inaccordance with its speeds and torques, which during the shift intervalare not sensibly varied until the torque paths have been reestablishedin the newly established ratio. y

The herein described features of novelty are believed to be especiallyadapted to the drive and to the control of large heavy land vehicles andsimilar apparatus, and provide special advantages in structure,arrangement, safety and convenience of control, not before revealedin'the prior art.

The foregoing demonstration of transmission I mechanism and controls isa complete description of a working device in accordance with thepresent invention, by way of example, and it is to be expresslyunderstood that the invention is not to r be limited strictly thereto,but that various changes in detail, design and in construction andarrangement may be made without departing from the spirit and scope ofthe invention as set forth in the appended claims, wherein I claim:

1. In power drives, an engine shaft and a load shaft, a variable speed.ratio gear unit coupling said shafts, said unit including input andoutput members coupled by gear elements with reactionmembers and havingplural clutch means energizable for individual or plural connectionbetween a fluid turbine-drive mechanism driving said unit and said loadshaft, an arrangement of said input and output members, said gearelements and reaction members adapted to provide reverse, low-directdrive and overspeed ratio drive, brake means for the said torquereaction members energizable for individual operation to establish bothforward and reverse drive of said load shaft, actuation means for saidclutch and said brake means, and controls for said actuation meansoperable in a shift pattern requiring the actuation of one clutch ofsaid clutch means with one brake of said brake means for an overspeedforward drive by said gear elements and the same brake with another ofsaid clutches for reverse drive by said gear elements and providingsimultaneous actuation of two clutches of said clutchmeans for low ratiodirect drive of said unit.

2. In the construction described in claim 1, the sub-combinaticn ofmanually-controlled operating means havinga coacting linkage with saidactuation means and with said controls effective to provide normalforward speed ratio changes within said shift pattern while maintainingunin terrupted torque between said engine and said load shafts.

3. In power transmitting variable speed ratio devices, a power shaft anda load shaft, a planetary gear unit coupling said shafts and adapted toprovide a plurality of forward overspeed ratios, direct drive andreverse drive, thru gearing having plural gear output members, saidgearing of said unit including an input member and with reaction membersadapted to establish reaction torque for transmitting drive thru thesaid gearing, said overspeed ratios being established by a pair ofactuatable .couplin clutches adapted to i connect said load shaft tosaid output members, a pair of actuatable reaction brakes for saidreaction members, actuating means for said forward overspeed clutchesand for said brakes, controls for said actuating means arranged forshift to reverse, to direct and to forward overspeed positions, andoperating means for said controls operable to provide said shiftpositioning 7 and operable such that for the controlled transition ofsaid controls between reverse drive and forward direct drive one of saidpair of brakes and one of said clutches of said pair of clutches arealternately engaged or released by manipulation of said operating means.

4. In power transmissions, an engine shaft, a load shaft and drivingmechanism coupling said shafts including a planetary gear unit providingdirect drive, two overspeed ratios and reverse drive, said unitconsisting of input and output members coupled by gearing includingreaction members and having actuatable clutches connecting said loadshaft with the output members of said unit and two brakes connected tostop the said two overspeed ratios, actuating means for I I saidclutches and brakes, control means for said actuating means, andoperating means for said control means connected to shift the same in apattern of connection effective to cause said actuating means to holdone of said clutches engaged while alternating the actuation of saidbrakes for establishing one or the other of said two overspeed ratios.

5. In power transmissions, an engine shaft, a load shaft, drivingmechanism coupling said shafts including a gearing unit providing directdrive, reverse drive and higher and lower overspeed ratios, said unitcomprising input and output members coupled by gear elements includingreaction members, a pair of clutches adapted to' connect said load shaftwith twoof said gear elements, brakes operable to stop said reactionelements of said unit, actuating means for said clutches and brakes,control means for said actuating means having stop positions fordirect-reverse and for said overspeed ratios, and operating mechanismfor said means operative such that the driving speed ratio of said unitmay be quickly changed from the said highest overspeed ratio to reverseby release of engagement of one of said clutches and actuation oftheother of said clutches, through the shifting of said control means fromone said overspeed ratio position to thereof and reaction brakesarranged to stop the reaction members thereof, fluid pressure actuatorsfor said clutches and brakes, fiuidpressure valving controlling theapplication ofpressure-to said actuators, including twovalves eachhaving end-point and central positions for determining the actuation ofsaid clutches and brakes in ac-,

cordance with a predetermined ratio shift pattern, and operatingmechanism for said valving which provides direct drive by said clutcheswith both valves in the same end-point position while all other ratiosare determined,with the valves in other than the same positions.

7. In power transmissions, an engine shaft and a load shaft, powertransmission mechanism coupling said shafts including a gearing unitproviding a plurality of overspeed ratios, direct and reverse drive,said unit consisting of a power input member, power output members andreaction.

19 members with gearing adapted to couple said shafts and having aplurality of output coupling clutches and reaction brakes operative uponsaid reaction members, said clutches andbrakes being selectivelyoperable to provide said overspeed ratios, direct and reverse drive,fluid pressure actuating means for said clutches and brakes, the

- said actuating means for said clutches being operable to engage atleast two of said coupling clutches alternatively, control valving forsaid means including means to supply fluid pressure through at least twopressure delivery passages, and operating mechanism for the positioningof said valving such that in direct drive the fluid pressure is admittedto the actuating means for both said clutches, and .that in reverse orfirst overspeed ratio the valving is positioned to admit fluid pressureto the clutch actuating means alternately,

8. In power drives, an engine shaft and a load shaft, a power traincoupling said shafts including a gearing unit consisting of power inputand output members connected by gear elements having reaction membersand having two ratio-determining reaction brakes operative upon thereaction members" thereof, and a plurality of clutches coupling saidgear elements to said load shaft, the arrangement providing reverse,direct and a first and a second overdrive ratio between said. shafts,actuation means operable by fluid pressure to energise said clutches andbrakes,

fluid pressure valving for said means, control means for said valvingeifective to establish fluid pressure in a predetermined shift controlpattern for sequential shift from reverse to direct, direct to firstoverdrive, and first overdrive to second overdrive by selectiveactuation of said clutches and brakes, and a supplementary valveoperable by said control means when placed in the direct drive selectingposition to cause vehicle braking as may be required in an emergency,established by actuation of both said brakes.

9.In power drives for vehicles, an engine, a finaldrive shaft, a fluidtorque converter driven by said engine and connected to transmit power Ito said shaft through a planetary gear unit driven by said torqueconverter and providing a plurality of speed ratios including reverse,direct and two overspeed drive ratios, said unit comprising power inputand output members potentially coupled by gear elements with reactionmembers and including a-flrst and a-second reaction brake operative uponthe reaction members of said unit and a first and a second couplingclutch connected to the gear elements thereof; actuation means for saidclutches and said brakes; and control means for said actuation meansproviding a reverse gear drive by requiring actuation of said flrstnamedbrake and said first named clutch, providing a direct drive by requiringactuation of both said coupling clutches, providing a first overspeedby" requiring actuation of said second named brake and said second namedclutch, and providing a second overspeed by requiring actuation of saidfirst named brake and said second named clutch.

10. In vehicle driving devices, an engine, a reduction gear drive drivenby said engine and driving a variable speed transmission assembly whichplanetary gears with torque-reaction sustainin and torque couplingmembers equipped with actuatable friction means for establishingselected drive ratios between said final drive and said converter unit,and manual ratio selection means for controlling the said actuatablefriction means of said overspeed unit effective to establish a pluralrange of forward driving speed ratios in said overspeed unit with thefluid torque converter operating at its maximum torque capacity.

11. In vehicle propelling mechanism, an engine providing a relativelyhigh torque over a driving range of predetermined low speeds, a.reduction gear drive driven by said engine and driving the impeller of afluid torque converter at reduced speeds to those of said engine andbeing operable at maximum efllciency over the said driving low speedrange of said engine, fixed reaction guide means and an output turbinemember included in said torque converter, a planetary gear unit drivenby the output member of said torque converter and driving a drivenshaft, said unit consisting of power input and output members coupled bygear elements with a plurality of reaction members, and provided with aplurality of actuatable friction members for stopping the rotation ofsaid reaction members and coupling the gear elements of said unit withsaid driven shaft, the arrangement being effective to establish aplurality of forward overspeed step ratios in said gear unit, a finaldrive reduction gearing driven by the driven shaft of said gear unit,and control means for selecting the actuation of said members in aplurality of actuation combinations thereof whereby the drive in theselected forward speed ratios proceeds at maximum torque capacity ofsaid fluid torque converter and under high torque of said engine.

12. In power control devices, an engine shaft and a load shaft, iflxedreduction gearing driven by said load shaft, variable speed ratiodriving mechanism coupling said shafts including a reduction gear drivenby said engine for driving a fluid torque converter providing drivingoutput shaft speeds of approximate one-to-six ratio to the speeds ofsaid engine, a step-ratio gear unit includes a primary fluid torque,converter unit, I

a secondary overspeed gear unit changeable to provide direct drive and aplurality of overspeed ratios with a flnal reduction drive connected tog the vehicle propelling means, the drive mechanism of said unitincluding input, output and reaction members coupled by constant ymeshed output of said, converter driven by said converter and drivingsaid flrstnamed reduction gearing comprising planetary gearingconsisting of power input and output members coupled by gear elementswith reactionsupporting members and having plural reactionsustainingbrakes and plural coupling clutches connecting said load shaft with gearelements of said unit and actuatable in a predetermined pattern toestablish reverse, direct, first overdrive and second overdrive speedratios between the and the input to said fixed reduction gearing,actuating means for said brakes and said clutches, shiftable controlmeans for said actuating means and an arrangement of said control meansoperative to establish reverse, direct, first overdrive and secondoverdrive ratios when said control means is'shifted sequentially.

13. In power transmissions, an engine shaft aasaosa 21 and a pluralityof coupling clutches for delivering torque to said output member, afinal reduction gear unit connected for power drive by said 22 shafts, avariable speed driving mechanism which includes the serial arrangementbetween said power and said load shafts of a primary fixed clutches toestablish the selected step speed ratios of said gearing unit, and ofvalves embodied in said control means arranged to be positioned in apredetermined pattern of ratio selection, involving the utilization ofthree effective positions for each of said valves.

15. In power transmissions, an engine, an engine shaft, a casingattached to said engine, a primary fixed reduction gear unit consistingof an annulus gear connected to said engine shaft, a sun gear attachedto the said casing,

with an output shaft, rotated by this arrangement at a fixed reducedspeed ratio to that of said engine shaft, a fluid torque converterassembly composed of an input impeller fastened to the output shaft ofsaid'primary gear' unit, a

' set of reaction blades fixed to said casing and an output rotor, thisassembly providing torque multiplication over the speed range providedby said primary unit from said engine; a variable speed gearing havingan input planetary carrier rotating with said rotor, compound sun andannulus gears with reaction brake drums, coupling clutches arranged toapply said brakes alternately for establishing one or the other of saidoverdrive speed ratios while one of said coupling clutches is engaged:

16. In power transmissions, an input shaft connected to a variable speedgearing unit which includes two coupling clutches for obtaining variablespeed ratios therein, said clutches having common external connectionbut each having an independently rotatable hub, an output shaft, a drumattached to said output shaft and enclosing said clutches, fluidpressure cylinders formed in said drum. a pair of annular pistons insaid cylinders for actuating said clutches under fluid pressure admittedto said cylinders, two fluid pressure passages formed in said drumleading to said cylinders, one for actuating each piston, a solid shaftdriven by an element of said gearing unit and connected to one ,of saidclutch hubs, a hollow shaft surrounding a portion of said solid shaftand connected to the other of said clutch hubs, a supporting sleeveexternal to said shafts and adjacent one end of said drum, fluidpressurefeed means in said sleeve and connected to said passages, and controlmeans external to said fluid feed means operative to direct fluidpressure selectively :or simultaneously to said passages for causingengagement of one orboth of said clutches or for causing release of bothclutches.

17. In power transmissions, power and load and a set of planet gearsmounted on a carrier integral ratio gear unit, a fluid torque converter,having an input impeller driven constantly by said unit and having anoutput rotor, a selective step ratio searing unit with to said rotor,said unit having a power output member coupled to the drive of saidinput mom-- her by a plurality of gear elements and including gearreaction and clutch coupling members the arrangement providing aplurality of forward overspeed ratios and reverse drive, and afinaldrlve fixed ratio unit driven by said output membet and drivingsaid load shaft. actuation means for determining the plurality of stepratios provided by said gearing unit, including brake means for saidreaction members and clutch means for said coupling members, andoperating controls for said actuation means characterized by a linkageerrangement which is effective to produce a range of forward drivinspeed ratios by said gearing unit while maintaining uninterrupted torquebetween said shafts,

18. In power transmissions, an engine, a speed control for said engine,an engine shaft and a load shaft, 9. fluid torque converter driven bysaid engine shaft and driving a variable speed gear ing unit embodying aplurality of reaction elements and providing reverse and a plurality offorward speed ratios, a drum adapted to be driven by said gearing unitand connected to drive said load shaft, a coupling clutch connectingsaid drum with a reaction element'of said gearing unit and arranged tobe actuated during reverse and during the lowest speed ratio forwarddrive, actuating means for said clutch, control means for saidtransmission efiective to cause actuation of said clutch when set forreverse drive by said unit and likewise when set for one of said forwardspeed ratios and operating means for said speed control and said controlmeans such that an operator by predetermined operation of said speedcontrol for said engine is enabled to" shift said control means andthereby said actuating means freely between forward and reverse withoutrelease of the said clutch during the transition interval between thesespeed ratios.

19. In power drive mechanisms for providing continuous torque over arange of driving speed ratios between a variable torque engine shaft anda load shaft, the combination of a fluid torque converter driven by thesaid engine shaft effective to transmit multiplied torque over a rangeof speeds and torques and having an output shaft coupled to the inputshaft of a variable speed ratio gear unit consisting of power input andoutput members coupled by gear elements with determined pattern toprovide a sequential selection of forward drives between the saidtransmission input shaft and the said load shaft consisting of direct, afirst overspeed and a second overspeed, the transitions between saidforward drive ratios being accomplished while maintaining continuoustorque between said engine and said load shafts, the said fluid torqueconverter automatically changing ratio in accordance with output torqueinversely to the changes in said variable speed ratio gear, such thatthe said transitions between direct and first oversoeed. or betweenflrst and second are achieved while the net a power input memberattached overall torque between the said engine-and said load shaftremains at a given value.

20. In power drive mechanisms, an arrangement of power devices connectedin series, consisting of a variable torque engine. a fluid torqueconverter providing automatic variations in multiplied torque inaccordance with the speed and load applied to the said converter, thesaid converter having an input impeller and an output rotor; and'avariable step ratio gearing unit driven by a fixed connection from saidrotor, said unit consisting of input and output members coupled by gearelements with torque-reaction sustaining members and withtorque-coupling clutch members, the arrangement thereof being adapted toyield changes of ratio obtained by friction actuators operable on saidreaction sustaining and torque coupling members to maintain torqueduring said ratio changes, the assembly driving a load shaftpower-operated means adapted to energise said friction actuators, andcontrol mechanism for said power-operated means providing apredetermined pattern for energising said friction actuators effectiveto establish said changes of ratio by said gearing resulting inautomatic 24 inverse changes of multiplied torque by said torqueconverter such that for a predetermined step ratio change in saidgearing, the overall torque between said engine and said load shaftremains approximately constant.

OLIVER K. KELLEY.

REFERENCES CITED The following references are of record in the file ofthis patent:

UNITED STATES PATENTS Number Name Date 2,235,370 Jandasek Mar. 18, 19412,100,191 Lapsley Nov. 23, 1937 2,134,398 Cotterman Oct. 25, 19382,144,795 Cotterman Jan. 24, 1939 2,205,794 Jandasek June 25, 19402,241,764 Bollinger- May 13, 1941 2,325,876 Pollard Aug. 3, 19432,221,393 Carnegie Nov. 12, 1940 1,609,782 Small Dec. 7, 1926 2,308,547Schneider Jan. 19, 1943 2,351,213 James June 13, 1944

